Multi-stage regulator for variable displacement pumps

ABSTRACT

Multi-stage regulator for lubricant pumps with continuously adjustable displaced volumes in particular for lubricant pumps, with a regulating piston to which pressure can be applied from pressure derived from the output side of the pump, and with increasing pressure can be displaced counter to the force of a flexible restoring element.

The present invention relates to a multi-stage regulator for pumps, inparticular for lubricant pumps, with a regulating piston to whichpressure can be applied from pressure derived from the output side ofthe pump, and with increasing pressure can be displaced counter to theforce of a flexible restoring element.

Such a regulator is, for example, known from regulatable vane-cellpumps. Generally in this case a piston with a circular cross-section isguided in a cylinder and pressure is applied to it by means of a branchconduit from the output side of the pump. A counter-compression springopposes resistance to the movement of the piston in the cylinderresulting from the action of the pump pressure. A drive pin for theadjustment ring of the vane-cell pump is located, for example, betweenthe piston and the spring. By displacement or pivoting of the adjustmentring relative to the rotor of the vane-cell pump, the eccentricity ofthe vane-cell pump is altered and thereby also the displaced volume ofthe individual cells of the pump as well as the hydraulic pumpingcapacity as a whole. The relative arrangement of piston, spring, rotor,adjustment ring and drive pin is in addition selected so that withrising pressure, as a result of the piston moving forward increasinglyin opposition to the force of the compression spring, the adjustmentring is pressed over the drive pin, in the sense of less eccentricityrelative to the rotor, so that the pumping capacity reduces and theoutput pressure sinks, or else is kept constant. The pump parameterswhich can be pre-determined, that is to say in concrete terms thegeometry and size of the rotor and adjustment ring, the maximum andminimum eccentricity thereof determined by the maximum regulatingdistance of the piston-spring-drive pin system, and the springcharacteristics, determine the properties of the regulator, that is tosay above all the maximum pressure which can occur at the pump output.The speed the load connected to the output side of the pump, whichrepresents a pressure sink with a more or less adjustable, and generallytemperature dependent, flow resistance.

Such pumps are often found in internal combustion engines, which aregenerally used over a wide range of revolution speeds and at variableoperating temperatures. In this case the pump shaft is normally directlyconnected to the crankshaft of the engine, and thus rotates at the samerevolution speed as the engine, or at least at a revolution speedproportional thereto.

If the oil pressure produced by such a pump is plotted (at a giventemperature) against the engine revolution speed, normally a more orless linear progression is found, up to a pre-determinable maximum valueof, for example 5 bar, wherein above the revolution speed at which thisvalue is reached, the pressure is kept constant with the aid of theregulator.

The oil pressure necessary for an internal combustion engine is alsodependent on the revolution speed, above all because the engine haspivot bearings and rotating parts and the lubricant has in part to beconveyed radially inwards from outside, counter to the effect ofcentrifugal force, to rotating parts in order to guarantee sufficientlubrication of such areas. As a result there is generally a non-linearincrease in the minimum necessary oil pressure with increasingrevolution speed.

The development of minimum oil pressure for a typical combustion engineis shown as curve a in FIG. 1, wherein at the same time the developmentof the oil pressure of a lubricating oil pump with pressure control,coupled to an engine, is also shown, as curve b.

The hydraulic capacity of a pump results from the product of the volumeflow produced and the pressure difference between the output and inputof the pump. Regulatable lubricant pumps such as, for example, vane-cellpumps, have the property of using pressure as the regulation parameter,as the pressure acts indirectly or directly on the adjusting memberwhich can alter the eccentricity of the piston ring, so that the amountof adjustment, that is to say that which can be altered by dealing withthe parameter of pressures on the pump, is the volume flow through thepump, as by altering the eccentricity of the adjustment ring thedisplaced volume is altered. Because of the construction of theregulatable vane-cell pump, at least over a wide part of the availablerevolution speed range, the pressure at the pump output is kept constantby adapting the displaced volume. As however the flow resistance, forexample in an engine, also alters depending on the operatingtemperature, generally the displaced volume necessary to maintain acertain pressure also alters at least depending on the temperature.Independently of the actual connection between the displaced volume andpressure, a load, for example an engine, has for a given operatingcondition a certain minimum lubricant requirement, and this is withrespect to both the lubricant pressure and the amount of lubricant.Conventional, regulated vane-cell pumps only regulate one of thesequantities, however, wherein the other quantity always adapts accordingto the given flow resistance. In addition, the pumps are generally setup for the maximum requirements of the engine with respect to the amountof lubricant and lubricant pressure. Therefore, with pressure controlsystems which are known in general for lubricant pumps and specially forvane-cell pumps, in particular in the mid-range of revolution speeds,the actual pressure delivered by the pumps and the amount of lubricantdisplaced is substantially greater than corresponds to the actualrequirement of the load, in concrete terms the internal combustionengine. The surplus hydraulic capacity of the pumps in this respectresults in their unnecessarily high consumption of operating energy.This means that despite a certain improvement compared to fixeddisplacement pumps, the displaced volume of which is simply partiallydiverted via pressure control valves, pressure controlled lubricantpumps still also do not work according to demand over wide areas.

Compared to this state of the art, the object of the present inventionis to provide a regulator for regulatable lubricant pumps, wherein thesimplest possible design makes possible a qualitative improvement to theregulation characteristics, which better corresponds to the actualrequirements of the engine with respect to pressure and displaced volumeof lubricant, and consequently operates largely without losses, that isto say without surplus hydraulic capacity.

This object is solved in that the regulating piston is composed of aplurality of part pistons to which pressure from the system can beapplied simultaneously, which are moveable over a limited regulatingdistance in a parallel manner, wherein a stop is provided for at leastone of the part pistons, which limits its regulating distance counter tothe direction of force of the restoring element when there is increasingpump output pressure, wherein at least one further part piston remainsmoveable counter to the force of the restoring element.

As long as the part pistons of which the regulating piston according tothe invention is composed move parallel as a common unit, such aregulator has the same qualitative pressure development as the regulatordescribed in the introduction for conventional, regulatable vane-cellpumps. That is to say, while with very low revolution speeds the initialpressure of the pump rises relatively steeply at first with increasingrevolution speed, without the control mechanism being actuated to anysubstantial extent, the part pistons, to which output pressure from thepump is applied, then move counter to the force of a restoring element,that is to say in general a compression spring, and thereby reduce thedisplaced volume of the pumps and consequently also their hydrauliccapacity. With this, in general a more or less constant pressure occursover a larger revolution speed range. Only when the revolution speedrises further and as a result there is a further displacement of thepart pistons, one of these part pistons makes contact with a stop andits further regulating distance is limited, and the force from the partpiston or part pistons which is still exerted against the restoringelement is not sufficient to effect a further displacement of this partpiston counter to the force of the restoring element, due to the smallersurface of the part piston or part pistons respectively, which are stillmoveable and upon which the control pressure acts. As the displacedvolume then also rises with increasing revolution speed, the outputpressure of the pump also increases until a second limit pressure isreached, at which point the force on the part piston or part pistonsstill moveable is sufficient for them to move counter to the force ofthe restoring element and thereby move a pusher, an adjusting member oranother element of the pump by means of which the displaced volumethereof, and thereby also the output pressure, is limited at a higherlevel compared to the first limit pressure.

The qualitative development of the resulting pressure curve over therevolution speed is shown as curve c in FIG. 2, wherein in FIG. 2 thedevelopment of curve b according to FIG. 1 is also shown again in dashedlines. As can be seen, curve c is substantially better "adapted" tocurve a than is curve b and the pump above all displaces the oil, withinan average revolution speed range, at a lower pressure and thereby alsoin smaller amounts (FIG. 3) which leads to a corresponding reduction inthe operating power needed for the pump.

In the preferred embodiment of the invention, a first part piston is anouter ring piston, while a second part piston is an inner centralpiston. This means the inner central piston is moveable within the outerring piston, wherein advantageously the movement is limited in at leastone direction, so that both part pistons move as one unit duringcorresponding operating conditions.

Advantageously, the central piston engages with a compression spring,serving as a restoring element, while the ring piston is provided with acatch by means of which the central piston is carried along counter tothe force of the compression spring when a corresponding pressure actsupon the pistons, so that the forces of the central piston and the ringpiston acting upon the compression spring are cumulative in total. In apreferred embodiment of the invention, the catch is provided in the formof a groove on the inside of the retaining ring embedded in the centralpiston.

The outer ring piston and the inner central piston can be provided withessentially independent guides such as, for example, guide bushes andthe like, however an embodiment of the invention is preferred in whichthe ring piston serves as the guide sleeve or cylinder for the centralpiston, so that in this way the external diameter of the central pistonis approximately equal to the internal diameter of the ring piston andboth pistons slide-fit tightly one within the other.

A concrete embodiment of the invention additionally provides that aguide cylinder is provided with a first cylindrical space for theguidance of the ring piston, so that in this way this cylindrical spacehas an internal diameter which corresponds approximately to the externaldiameter of the ring piston, whereby this first cylindrical spacenarrows in a step-wise manner and in this way adjoins a secondcylindrical space, the internal diameter of which is at least as largeas the external diameter of the central piston, wherein advantageouslythe axes of the central piston, ring piston, and the two cylindricalspaces of the guide cylinder should also be aligned with one another. Itis not necessary, however, for the second cylindrical space to form aguide for the central piston, at least not when it is already tightlyinserted in the ring piston.

In this embodiment the step-wise narrowing of the first cylindricalspace with respect to the second cylindrical space acts as a stop forthe ring piston, while with a suitable application of pressure, thecentral piston can move further into the second cylindrical spacecounter to the effect of a compression spring, wherein said space doesnot necessarily have to be cylindrical.

The multi-stage regulator according to the invention can naturally alsobe provided with additional ring pistons between the (outer) ring pistonand the central piston, which successively engage, following the outerpiston, with respectively assigned stops, and the central piston can forits part also be configured as a hollow cylinder.

An embodiment of the invention is particularly preferred in which thestop for at least one part piston, preferably for the ring pistondescribed, is variable, and is preferably variable dependent upon thetemperature, such that at lower temperatures a greater regulatingdistance is available for the part pistons being stopped than at highertemperatures.

With a fixed stop and when the lubricant is at a low temperature, ahigher pressure builds up at relatively low revolution speeds, whichmeans that the first part piston comes to a stop relatively early and asa result the displaced volume or the displacement pressure increasesprematurely from the first pressure control limit to the higher level atrelatively low revolution speeds. On the other hand with a variablestop, it is possible to delay the stopping time, above all at lowlubricant temperatures, or respectively to adjust it at excessively highrevolution speeds.

It is naturally also possible to configure the multi-stage regulatoraccording to the invention not only in the form of ring pistons packedone inside another, but instead corresponding pistons can be arrangedadjacently, wherein initially they exert a force in common upon theadjustment ring and after a pre-determinable regulating distance one ormore of the pistons comes into contact with a stop.

Further advantages, features and possibilities for application of thepresent invention will be made clear by the following description of apreferred embodiment, and the drawings with reference thereto. There isshown:

FIG. 1 the development of pressure over engine speed with a conventionalregulated vane-cell pump, and a typical minimum oil pressure for anengine, depending on the revolution speed,

FIG. 2 a representation of oil pressure curves corresponding to FIG. 1,wherein additionally the regulation characteristic for the pressurecontroller according to the invention is shown,

FIG. 3 the mass rate of flow of the engine, resulting from curve caccording to FIG. 2, depending on the revolution speed for differentconstant and variable pressures,

FIG. 4 a section through a vane-cell pump with the pressure controlapparatus according to the invention,

FIG. 5 the regulation characteristic of a multi-stage regulator with afixed stop, at different temperatures,

FIG. 6 a regulation characteristic of the multi-stage regulatoraccording to the invention, optimised by a temperature-dependentvariable stop, and

FIG. 7 the regulating part of a vane-cell pump with a variable stop.

FIG. 8 is a sectional view of an alternate embodiment of a multi-stageregulator according to the invention wherein pistons are arrangedadjacently.

FIG. 1 shows the development of two oil pressure curves a and b, for aninternal combustion engine, over engine speed. With this, curve a showsthe minimum oil pressure required by the engine at the respectiverevolution speed for maintaining a reliable lubrication of all theengine parts, while curve b shows the development of pressure which isactually produced by a regulated lubricant pump, in concrete terms aregulated vane-cell pump. In this case, the steepness of the initialrise in curve b is also dependent on the current operating temperature,and respectively from the flow resistance of the oil consuming pointsand incoming supplies. The disproportional rise in curve a withincreasing revolution speed is connected to the fact that bearings andother rotating parts have to be supplied from the outside with lubricantand that with this the lubricant or oil has partly to be pressedradially inwards counter to the effect of centrifugal force, wherein thecentrifugal force increases in a quadratic manner with the revolutionspeed.

The development of curve b results above all from the fact that thedisplaced volume of a vane-cell pump is approximately proportional tothe revolution speed, so that with constant flow resistance the pressurefirstly rises relatively fast with increasing revolution speed, and atthe same time, however, the regulating mechanism, composed of a pistonand a spring, downwardly regulates the displaced volume and then limitsit such that a substantially constant output pressure occurs above acertain revolution speed. This constant output pressure must, however,correspond at least to the pressure which is required by the engine atthe highest possible revolution speed and the highest operatingtemperature.

As, on the other hand, the maximum displaced volume geometry of alubricant pump has to be set up with respect to the two most criticalpoints of operation, namely the minimum revolution speed and the maximumoperating temperature, in general the maximum output pressure is reachedat far below the highest revolution speed, even with hot engine oil.

FIG. 1 shows that in particular in the mid-range of revolution speeds,the pressure produced by the pump is substantially greater than theminimum oil pressure required by the engine. At the same time, becauseof this higher pressure a larger amount of oil is drawn, wherein thehydraulic pump capacity is proportional to the product of pressure andamount of oil drawn per unit time. If therefore the pressure and amountof oil drawn are respectively double that of an oil pressure set up forthe minimum requirement, the hydraulic capacity required for this isfour times the minimum capacity required.

It is also shown that from the point of view of a lower powerconsumption by the pump, and consequently also a reduction in the energyconsumption of the engine it would be advantageous, particularly in themid-range of revolution speeds, to approximate the oil pressure producedby the pump to the minimum oil pressure required for the engine. This isdone with a regulator according to the present invention which,represented in this case by two two-stage controls, produces a pressuredevelopment as shown qualitatively in FIG. 2 by curve c. For bettercomparison, a curve b, corresponding to the curve b in ˜FIG. 1, is alsoshown in dashed lines in FIG. 2.

FIG. 2 shows that particularly in the range in which the distance apartbetween the curves a and b is greatest, and consequently the pump isproducing a particularly excessive amount of hydraulic pump capacity,and consuming a corresponding amount of driving power, the pressuredevelopment according to curve c decreases significantly and the curve ais very much closer, so that the pump also consumes correspondingly lesspower when its output pressure is set up according to curve c.

In FIG. 3, the mass rates of flow of lubricant are plotted over enginespeed in several curves. Curves a' and b' respectively show the massrates of flow of the engine, which are obtained when the pressure isalways kept at a constant value of 2.5 bar or 5 bar respectively (whichcorresponds to the limit pressures of curve c in FIG. 2). As can be seenfrom FIG. 2, these are approximately the two pressure stages at whichthe pressure regulator according to the invention keeps the outputpressure according to curve c of a vane pump constant over a certainrange of revolution speeds. The actual mass flow rate of the engine whenthe pressure is regulated according to curve c in FIG. 2 is then shownin the form of curve c' in FIG. 3. It is clear that the mass flow ratedevelopment in the ranges where the pressure is kept constant alsofollows the isobars in FIG. 3. However, it can also be seen from FIG. 2that a constant pressure of 2.5 bar would not be sufficient to reliablylubricate the engine at high revolution speeds, so that the change-overto a higher pressure and to a higher mass flow rate is necessary and,according to curves c and c' in FIGS. 2 and 3, also occurs.

The development of curves c and c' in FIGS. 2 and 3 is obtained by avane-cell pump with a pressure regulator, shown by way of example, asrepresented schematically in FIG. 4.

FIG. 4 shows a vane-cell pump 10 with a housing 16, a rotor 12, vanes 14driven by the rotor and an adjustment ring 11. The adjustment ring 11has radial inlet apertures 13, through which the cell volume formedbetween the vanes 14 and the adjustment ring 11, which, with thedirection of rotation shown, increases in the upper part of theadjustment ring in the direction of rotation, sucks in oil from thesuction space 15. In the lower part of the adjustment ring where thecell volume increasingly reduces in the direction of rotation, the oilis pressed out through output apertures which are not shown, or are notvisible. The displaced volume per rotation of the rotor is dependent onthe difference in the changes in volume of the individual cells in thetop and the bottom area of the adjustment ring, which again aredetermined by the eccentricity of the adjustment ring 11 with respect tothe rotor 12. The adjustment ring 11 is pivotably mounted and isprovided on its lower end, distanced from the swing shaft 17, with anadjustment pin 6, which engages on the one hand with a regulating piston1 and on the other hand with a compression spring 2. The compressionspring 2 is contained in a recess in the housing 16, and is supported byan embedded sleeve 8. The regulating piston 1 is composed of a centralpiston 3 which engages with the pin 6, and consequently acts directlycounter to the force of the spring 2, and a ring piston 4, closelyguiding the central piston 3. The ring piston 4 and the central piston 3are also contained in a recess in the housing 16, wherein the ringpiston 4 is guided in a pressure sealed but slidable manner in therecess in the housing which contains it. An end stopper 7 retains thepiston 1 in the housing 16 and at the same time produces a pressuresupply aperture 19 from which the output pressure is applied to thecentral piston 3 and the ring piston 4. The output pressure or systempressure p, produced by the pump, thus prevails in the space 18 betweenthe stopper 7 and the regulating piston 1. This acts on both the endface of the central piston 3 and the end face of the ring piston 4. Inthe internal wall of the ring piston 4, a continuous groove is providedin which a sealing ring 5 is contained, which ensures that when the ringpiston 4 moves to the left according to FIG. 4, it collides with theretaining ring on the end face of the central piston 3, and is obligedto carry said piston along for as far as the force of opposition exertedby the spring 2 allows. As long as the retaining ring 5 of the ringpiston 4 is non-positively engaged with the end face of the centralpiston 3, the two part pistons 3, 4 act as a unitary piston whichpresses against the pin 6 and thereby against the compression spring 2with a force which corresponds to the product of the prevailing pressuremultiplied by the total end surface of the two part pistons 3, 4.

As the displaced volume of the pump reacts relatively sensitively tosmall adjustments to the adjustment ring 11 or respectively to theregulating pin 6, a regulating mechanism is obtained which produces asubstantially constant pressure depending on the piston surface of thepistons 3, 4 and the spring constants of the compression spring 2 over acertain range of revolution speeds of the pump, which, according to FIG.2, can be approximately 2.5 bar. With increasing revolution speed,however, the front end face of the ring piston 4 then reaches a stop 9,which is formed by a step-wise transition of the cylindrical guide spacefor the ring piston 4 to a cylindrical space with a smaller diameter inwhich the central piston 3 can still move with clearance. If therevolution speed now increases further, and with this the pressure, thering piston 4 can no longer participate in a displacement of theadjusting pin 6, as the force of pressure acting on its right-hand endface is captured by the stop 9 and can no longer be transferred to thecentral piston 3. With increasing revolution speed, at this stage thepressure will further increase until at last the force of pressureacting solely on the end face of the central piston 3 is sufficient toovercome the opposing force of the spring 2, whereupon the pressure isagain regulated to a substantially constant level, albeit now at ahigher level of approximately 5 bar (see FIG. 2).

FIG. 5 shows curve c according to FIG. 2, for three differenttemperatures of the lubricant, that is for 30° C., 80° C. and 130° C.The curves are preferably to be interpreted as follows, with referenceto the control mechanism in FIGS. 4 and 7 respectively. In the case ofthe curve shown by a dashed line, for a lubricant temperature ofapproximately 30° C., at very low revolution speeds there is firstly acontinuous increase in the oil pressure until the regulator actuallystarts operating. With increasing revolution speed, the displaced volumeand pressure increase, so that with a pressure of approximately 2.5 bar,the control mechanism begins to operate, in that both pistons move thepiston ring 11 counter to the action of a restoring element such as thespring 2 in FIG. 4, and thereby keep the pressure constant. Finally, theouter ring piston collides with the stop 9, so that the force now actingsolely on the part piston 3 is insufficient to further compress thespring 2. In this way, the regulator is firstly inactive again, so thatthe pressure and volume flow can again increase with the revolutionspeed, which is shown diagrammatically in FIG. 5 by the increase from2.5 bar to 5 bar. The moment of contact of the ring piston 4 with thestop 9 is thus characterised by the kink in the curve of oil pressureover the engine speed at the end of the 2.5 bar level. When the pressureof 5 bar is reached, the force acting upon the part piston 3 is thenalso sufficient to push the adjustment ring 11 counter to the action ofthe spring 2, so that the pressure is then kept at the constant level of5 bar up to the maximum revolution speed.

At higher temperatures, the lubricant flows more easily through thelines and other flow resistors, such as friction bearings, the result ofwhich is that the development of the oil pressure over the revolutionspeed is qualitatively equal, however in quantitative terms indicates aless steep increase and a displacement of the change-over point to thehigher speed of revolutions. This means that the two part pistons firststart to move at a higher revolution speed, namely when an intermediatelevel of 2.5 bar is reached and keep the pressure substantially constantbecause of their inward movement until the outer ring piston 4 againcollides with the stop 9 and a renewed increase in pressure toapproximately 5 bar occurs. As shown in FIG. 5, the development of thecurves at 30° C. and 130° C. is increasingly less steep, and thechange-over point, and also in particular the stopping of the outer ringpiston 4 on the corresponding stop 9, is displaced to higher revolutionspeeds with rising temperature. By means of a temperature-dependentstop, it is however possible that at least the stop of the outer ringpiston also moves at low temperatures to lead to higher revolutionspeeds, that is to say gets nearer to an ideal curve or respectively acurve engine requirements. This development is shown in FIG. 6 and isproduced by a temperature-dependent stop 9' as shown in a regulatoraccording to FIG. 7. The temperature-dependent stop 9' can, for example,be an expansion element, such as fitted in thermostatic valves, whereina special adjustment has to take place for the new purpose, but theprinciple of operation is however essentially the same. In concreteterms, the temperature-dependent stop 9' operates so that at lowertemperatures the stopping point for the part piston 4 moves to the left,in other words the expansion elements shrinks or contracts. In this way,at lower lubricant temperatures, which because of close contact alsodetermine the temperature of the expansion element, that is to say thetemperature-dependent stop 9', the part piston 4 has a greater availableregulating distance, which in concrete terms means that the secondchange-over point, that is to say the contact of the part piston 4 onthe stop 9', takes place at a higher revolution speed compared to FIG.5. With increasing temperature, the expansion element expands and thestop 9' is moved further to the right and thereby shortens the availableregulating distance for the part piston 4. In this way, in contrast toFIG. 5, the beginning of the increase from the 2.5 bar level to the 5bar level no longer moves at higher revolution speeds, despite thedecreasing flow resistance in the system, or in other words, at lowtemperatures this change-over point has already been moved so close tothe requirements curve that with increased temperature, a furthermovement of this point nearer to curve a is not desired. In this way therelatively advantageous case shown in FIG. 6, wherein the stopping ofthe outer ring piston 4 is largely independent of the temperature, canbe obtained, that is to say it always takes place at approximately thesame revolution speed at any temperature. It can be seen that comparedto FIG. 5, in this way a still more advantageous approximation to therequirements of the engine according to curve a is obtained even atlower temperatures, while the development at high temperatures remainssubstantially unaffected.

FIG. 8 shows an alternate embodiment wherein two separate pistons, afirst piston 3 and a second piston 4, are movable in two parallel,cylindrical bores 1a, 1b which extend from a common pressure chamber 18.The second piston 4 is provided with a lower lateral extension engagingthe bottom of the first piston 3 while the first piston 3, on theopposite end thereof, is acting against an adjustment pin 6 which inturn is urged to the opposite direction by a helical spring 2. A stop 9is provided for the second piston 4 and it may easily be recognized thatupon increasing pressure, both first and second pistons 3, 4 will bemoved together against the action of the spring 2 until the secondpiston 4 reaches the stop 9, whereupon also the movement of the firstpiston 3 is stopped because the force exerted by the first piston 3alone may not be sufficient to overcome the counteracting force ofhelical spring 2. However, once pressure P is further increased by asufficient amount, the first piston 3 again starts to move against theaction of the spring 2 (possibly until a final stop is reached for thefirst piston 3, too). The resulting oil pressure curves and curves andengine flow rates, each versus the engine revolutions in a correspondinglubrication system, will then be the curves c and c' as shown in FIGS. 2and 3, respectively, just as for the concentrical pistons shown in FIGS.4 and 7.

We claim:
 1. Multi-stage regulator for a variable displacement volumepump with a pressure regulating piston to which a pressure (P) derivedfrom the output side of a said variable displacement volume pump can beapplied, and which upon increasing pressure can be displaced counter tothe force of a flexible restoring element, said regulator comprising aregulating piston having a plurality of part pistons, including at leasta first part piston and a second part piston, to which pressure can besimultaneously supplied, said part pistons being engagable with eachother such that they can be moved simultaneously over a limitedregulating distance in a parallel manner, wherein for at least saidsecond part piston of said plurality of part pistons a stop is providedwhich limits the regulating distance of said second part piston inopposition to the direction of force of said restoring element as thepressure increases, wherein said first part piston of said plurality ofpart pistons still remains movable counter to the force of the restoringelement, wherein said first and second part pistons have transferringmeans therebetween such that the force acting on said second part pistonwhen being in engagement with said first part piston is transferred tosaid first piston via said second part piston.
 2. Multi-stage regulatoraccording to claim 1, wherein the part pistons are parallel pistonswhich are not packed on another.
 3. Multi-stage regulator according toclaim 1, wherein said first part piston comprises an inner centralpiston and wherein said second part piston comprises an outer ringpiston.
 4. Multi-stage regulator according to claim 3, wherein thecentral piston engages with the restoring element wherein said restoringelement comprises a compression spring, and wherein the ring piston isprovided with said transferring means comprising a catch for carryingalong the central piston.
 5. Multi-stage regulator according to claim 4,wherein said catch comprises a locking ring.
 6. Multi-stage regulatoraccording to any one of claims 3 or 4, wherein the external diameter ofthe central piston is approximately the same as the internal diameter ofthe ring piston, so that the central piston can be received in a tightsliding fit in the ring piston.
 7. Multi-stage regulator according toany one of claims 3 or 4, further comprising a guide cylinder having afirst cylinder space, the internal diameter of which correspondsapproximately to the external diameter of the ring piston, whereinadjoining by means of a step-wise narrowing of the first cylindricalspace is a second cylindrical space, the internal diameter of whichcorresponds to the external diameter of the central piston, wherein thetransitional step between the cylindrical spaces forms a stop for thering piston.
 8. Multi-stage regulator according to any one of claims 3,4 or 1, wherein the stop is adjustable.
 9. Multi-stage regulatoraccording to claim 8, wherein the stop is automatically adjustabledependent upon the temperature.
 10. Multi-stage regulator according toclaim 9, wherein the direction of adjustment is selected so that atlower temperatures, a greater regulating distance is produced for thepart piston being stopped.
 11. Multi-stage regulator for a variabledisplacement volume pump with a pressure regulating piston to which apressure (P) derived from the output side of a said variabledisplacement volume pump can be applied, and which upon increasingpressure can be displaced counter to the force of a flexible restoringelement, said regulator comprising a regulating piston having aplurality of part pistons, including at least a first part piston and asecond part piston, to which pressure can be simultaneously supplied,said part pistons being engagable with each other such that they can bemoved simultaneously over a limited regulating distance in a parallelmanner, wherein for at least said second part piston of said pluralityof part pistons a stop is provided which limits the regulating distanceof said second part piston in opposition to the direction of force ofsaid restoring element as the pressure increases, wherein said firstpart piston of said plurality of part pistons still remains movablecounter to the force of the restoring element; wherein said first partpiston comprises an inner central piston and wherein said second partpiston comprises an outer ring piston.
 12. Multi-stage regulatoraccording to claim 11, wherein the central piston engages with therestoring element, wherein said restoring element comprises acompression spring, and wherein the ring piston is provided with a catchfor carrying along the central piston.
 13. Multi-stage regulatoraccording to any one of claims 11 or 12, wherein the external diameterof the central piston is approximately the same as the internal diameterof the ring piston, so that the central piston can be received in atight sliding fit in the ring piston.
 14. Multi-stage regulatoraccording to any one of claims 11 or 12, further comprising a guidecylinder having a first cylinder space, the internal diameter of whichcorresponds approximately to the external diameter of the ring piston,wherein adjoining by means of a step-wise narrowing of the firstcylindrical space is a second cylindrical space, the internal diameterof which corresponds to the external diameter of the central piston,wherein the transitional step between the cylindrical spaces forms astop for the ring piston.
 15. Multi-stage regulator according to claim12, wherein the stop is adjustable.
 16. Multi-stage regulator accordingto claim 15, wherein the stop is automatically adjustable dependent uponthe temperature.
 17. Multi-stage regulator according to claim 16,wherein the direction of adjustment is selected so that at lowertemperatures, a greater regulating distance is produced for the partpiston being stopped.